Balance device for engines

ABSTRACT

A balancer device for engines, in which first and second balancer shafts are arranged in parallel to a crankshaft and the crankshaft rotatingly drives first and second balancer weights at the same speed as that of the crankshaft, characterized in that the first balancer weight is arranged on one end of the first balancer shaft in a direction along the crankshaft, the second balancer weight is arranged on the other end of the second balancer shaft in the direction along the crankshaft, and the first and second balancer shafts are made close to the crankshaft so that loci of rotation of the first and second balancer weights partially overlap a locus of rotation of that portion of the crankshaft, which is coupled to a conrod when viewed in the direction along the crankshaft.

TECHNICAL FIELD

The present invention relates to a balancer device for engines, comprising two balancer shafts.

BACKGROUND ART

In order to prevent generation of uncomfortable vibrations in an engine, a balancer device for engines is in some cases adopted, in which first and second balancer shafts are arranged in parallel to a crankshaft and the balancer shafts are reversely rotated at the same speed as that of the crankshaft.

With the conventional two-axis type balancer devices, there are some cases, in which surroundings of a crankcase become large in size and hindrance is caused in weight balance of a vehicle according to the arrangement and structure of balancer shafts, and there is a fear of causing problems, such as structural complication, an increase in the number of parts, degradation in assembling quality, loss in engine output, or the like.

The invention has been thought of in view of the above conventional problems, and has its object to provide a balancer device for engines, which can prevent large-sizing of surroundings of a crankcase, eliminates hindrance in weight balance, and can avoid problems, such as structural complication, an increase in the number of parts, degradation in assembling quality, loss in engine output, or the like.

DISCLOSURE OF THE INVENTION

The invention provides a balancer device for engines, in which first and second balancer shafts are arranged in parallel to a crankshaft and the crankshaft rotatingly drives first and second balancer weights at the same speed as that of the crankshaft. The first balancer weight is arranged on one end of the first balancer shaft in a direction along the crankshaft The second balancer weight is arranged on the other end of the second balancer shaft in the direction along the crankshaft. The first and second balancer shafts are made close to the crankshaft so that loci of rotation of the first and second balancer weights partially overlap a locus of rotation of that portion of the crankshaft, which is coupled to a conrod when viewed in the direction along the crankshaft.

The invention includes a feature that both sides of that portion of the crankshaft, which is coupled to the conrod, are born through first and second journal bearings by side walls of a crank chamber. The first and second balancer weights are rotatably supported by the first and second balancer shafts, and balancer driven gears formed on the first and second balancer weights mesh with crank drive gears, which are arranged close to and outside the first and second journal bearings.

The invention provides a balancer device for engines, in which first and second balancer shafts are arranged in parallel to a crankshaft and the crankshaft rotatingly drives first and second balancer weights at the same speed as that of the crankshaft. The first and second balancer shafts serve as coupling bolts for a left and right split type crankcase, and the first and second balancer weights are rotatably supported by the first and second balancer shafts.

The invention includes a feature that the first balancer weight is arranged on one end of the first balancer shaft in a direction along the crankshaft. The second balancer weight is arranged on the other end of the second balancer shaft in the direction along the crankshaft. The first and second balancer shafts are made close to the crankshaft so that loci of rotation of the first and second balancer weights partially overlap a locus of rotation of that portion of the crankshaft, which is coupled to a conrod when viewed in the direction along the crankshaft.

The invention includes a feature that both sides of that portion of the crankshaft, which is coupled to the conrod, are born through first and second journal bearings by side walls of a crank chamber. The balancer driven gears formed on the first and second balancer weights mesh with crank drive gears, which are arranged close to and outside the first and second journal bearings.

The invention includes a feature that the first and second balancer shafts comprise weight supports for supporting the first and second balancer weights, and crankcase-interior portions are arranged adjacent to the crankshaft. The crankcase-interior portions are formed to have a smaller diameter than that of the weight supports.

The invention includes a feature that oil introduction passages are formed on the first and second balancer shafts to communicate the crank chamber and balancer-weight bearing portions to each other.

The invention includes a feature that axes of rotation of the first and second balancer driven gears are made eccentric from axes of the first and second balancer shafts, and the first and second balancer shafts are rotated to thereby enable regulating backlash between the first and second balancer driven gears and the first and second crank drive gears.

The invention includes a feature that backlash regulation mechanisms are provided on the same end sides of the first and second balancer shafts in the direction along the crankshaft.

The invention includes a feature that the first and second crank drive gears comprise a ring-shaped gear portion arranged around a disk-shaped base portion fixed to the crankshaft to be rotatable relative thereto and have teeth to mesh with the balancer driven gears, and U-shaped buffer springs are interposed between the gear portion and the base portion.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a right side view showing an engine according to an embodiment of the invention.

FIG. 2 is a planar development of a cross section of the engine.

FIG. 3 is a left side view showing a valve gear of the engine.

FIG. 4 is a back surface view showing a cross section of the valve gear.

FIG. 5 is a planar development of a cross section showing a balancer device of the engine.

FIG. 6 is a bottom view showing a cylinder head of the engine.

FIG. 7 is a bottom view showing a cylinder body of the engine.

FIG. 8 is a cross sectional side view showing a connection of the cylinder head and the cylinder body of the engine.

FIG. 9 is across sectional side view showing a connection of the cylinder head and the cylinder body of the engine.

FIG. 10 is across sectional side view showing a connection of the cylinder body and a crankcase of the engine.

FIG. 11 is a left side view showing a balancer device of the engine.

FIG. 12 is an enlarged, cross sectional view showing a mount for a holding lever of the balancer device.

FIG. 13 is a side view showing constituent parts of a turning lever of the balancer device.

FIG. 14 is a side view showing a buffer structure of a balancer drive gear of the balancer device.

FIG. 15 is a right side view showing the balancer device.

FIG. 16 is a cross sectional, right side view showing a bearing bracket of the engine.

FIG. 17 is a cross sectional, left side view showing a bearing bracket of the engine.

FIG. 18 is a schematic view showing a configuration of a lubricating device of the engine.

FIG. 19 is a view showing a configuration of the lubricating device.

FIG. 20 is a cross-sectional side view showing surroundings of a lubricating oil pump of the lubricating device.

FIG. 21 is a cross sectional, left side view showing the lubricating device.

FIG. 22 is a cross sectional side view showing a balancer shaft for illustration of an embodiment of the invention of claim 7.

BEST MODE FOR CARRYING OUT THE INVENTION

An embodiment of the invention will be described below with reference to the accompanying drawings.

FIGS. 1 to 21 illustrate a balancer device for engines, according to an embodiment of the invention.

In the drawings, the reference numeral 1 denotes a water-cooling type 4-cycle single-cylinder five-valve engine, the engine having a general construction, in which a cylinder body 3, a cylinder head 4, and a head cover 5 are stacked and clamped to one another on a crankcase 2 and a piston 6 arranged slidably within a cylinder bore 3 a of the cylinder body 3 is connected to a crankshaft 8 through a conrod 7.

The cylinder body 3 and the crankcase 2 are coupled to each other by screwing four case bolts 30 a, which extend through a lower flange (case-side flange) 3 b, into a cylinder-side mating surface 2 e of the crankcase 2. More concretely, the case bolts 30 a are threaded into bolt-coupling portions (coupling boss portions) 12 c of iron-alloy bearing brackets 12, 12′ (described later), which are embedded into left and right walls of the crankcase 2 of aluminum alloy by insert-casting. In addition, the reference numeral 31 a denotes dwell pins for positioning of the crankcase 2 and the cylinder body 3.

Also, the cylinder body 3 and the cylinder head 4 are coupled to each other by means of two short head bolts 30 b and four lengthy head bolts 30 c. The short head bolts 30 b are threadedly implanted into a lower portion of an intake port 4 c of the cylinder head 4 and a lower portion of a discharge port to extend downward to extend through an upper flange 3 f of the cylinder block 3 to project downward. By screwing cap nuts 32 a onto lower projecting portions of the short head bolts 30 b, the upper flange 3 f and hence the cylinder body 3 are clamped to a cylinder-side mating surface 4 a of the cylinder head 4.

Also, the lengthy head bolts 30 c are threadedly implanted into the lower flange 3 b of the cylinder body 3 to extend upward to extend through a flange 4 b of the cylinder head 4 from the upper flange 3 f of the cylinder body 3 to project upward. By screwing cap nuts 32 b onto upper projecting portions of the lengthy head bolts 30 c, the lower flange 3 b and hence the cylinder body 3 are clamped to the cylinder-side mating surface 4 a of the cylinder head 4.

In this manner, at the time of coupling the cylinder body 3 and the cylinder head 4 to each other, not only the upper flange 3 f of the cylinder body 3 is clamped and fixed to the cylinder head 4 by means of the short head bolts 30 b and the cap nuts 32 a, but also the lengthy head bolts 30 c are implanted into the lower flange 3 b, which is bolted and coupled to the mating surface 2 e of the crankcase 2, and the lengthy head bolts 30 c and the cap nuts 32 b are used to clamp and fix the cylinder body 3 to the flange 4 b of the cylinder head 4, whereby the cylinder body 3 and the four lengthy head bolts 30 c bear a tension load produced by combustion pressure and it is possible to correspondingly reduce a load acting on the cylinder body 3. As a result, it is possible to decrease stresses generated especially in an axially intermediate portion of the cylinder body 3, thus enabling ensuring durability even in the case where the cylinder body 3 is decreased in wall thickness.

Incidentally, in the case where only the upper flange 3 f of the cylinder body 3 is coupled to the cylinder head 4, an excessive tension stress is generated in the axially intermediate portion of the cylinder body 3, and in extreme cases, there is a fear of cracks being generated in the portion, while according to the embodiment the presence of the lengthy head bolts 30 c can avoid generation of such excessive tension stress in the intermediate portion of the cylinder body, thus enabling preventing generation of cracks.

Also, at the time of implanting the lengthy head bolts 30 c into the lower flange 3 b, the lengthy head bolts 30 c are arranged in the vicinity of the case bolts 30 a for clamping of the crankcase, whereby the load produced by combustion pressure can be surely transmitted to the crankcase 2 through the lengthy head bolts 30 c and the cylinder body 3 from the cylinder head 4, which enables improving durability against the load.

Here, the right bearing bracket 12′ comprises a boss portion 12 b, into a bearing hole 12 a of which a right bearing 11 a′ for the crankshaft 8 is press-fitted and inserted, as shown in FIGS. 5 and 16. The bolt-coupling portions 12 c, 12 c extend upward to a neighborhood of the cylinder-side mating surface 2 e of the crankcase 2 from front and back portions of the boss portion 12 b with the crankshaft 8 therebetween, when viewed in a direction along the crankshaft 8.

Also, the bolt-coupling portions 12 c, 12 c extend upward to a neighborhood of the cylinder-side mating surface 2 e of the crankcase 2 from front and back portions of the left bearing bracket 12 with the crankshaft 8 therebetween, when viewed in the direction along the crankshaft 8, as shown in FIGS. 5 and 17. Also, the boss portion 12 b is formed with a collar hole 12 e, into which a bearing collar 12 d of iron having a larger outside diameter than a balancer drive gear 25 a described later is press-fitted. The left crankshaft bearing 11 a is fitted and inserted into the bearing hole 12 a of the bearing collar 12 d.

Here, the bearing collar 12 d serves to enable the crankshaft 8 to be assembled into the crankcase 2 in a state, in which a gear body 25 having the balancer drive gear 25 a is press-fitted onto the crankshaft 8.

Also, as shown in FIG. 5, a seal plate 25 d is interposed between the gear body 25 and the bearing (first journal bearing) 11 a on a left shaft portion 8 c of the crankshaft 8. An inside diameter portion of the seal plate 25 d is interposed between the gear body 25 and an inner race of the bearing 11 a, and a slight clearance is present between an outside diameter portion of the seal plate and an outer race of the bearing 11 a to avoid interference therebetween. Also, an outer peripheral surface of the seal plate 25 d is in sliding contact with an inner peripheral surface of a flange 12 h of the bearing collar 12 d.

Furthermore, a seal cylinder 17 i is interposed between the bearing (second journal bearing) 11 a′ and a cover plate 17 g on a right shaft portion 8 c′ of the crankshaft 8. An inner peripheral surface of the seal cylinder 17 i is fitted onto and fixed to the right shaft portion 8 c′. Also, an outer peripheral surface of the seal cylinder 17 i is formed with a seal groove of a labyrinth structure, and in sliding contact with an inner surface of a seal hole 2 p formed in a right case portion 2 b.

In this manner, the seal plate 25 d and the seal cylinder 17 i are present outside the bearings 11 a, 11 a′ on the left and right shaft portions 8 c, 8 c′ of the crankshaft 8 whereby leakage of pressure from a crank chamber 2 c is prevented.

In this manner, according to the embodiment, the bolt-coupling portions (coupling boss portions) 12 c, 12 c extending toward the cylinder body 3 are integrally formed on both sides of the iron-alloy bearing members 12, 12′ forbearing of the crankshaft, cast into the crankcase 2 of aluminum alloy to interpose therebetween a cylinder bore axis A, and the case bolts 30 a for coupling the cylinder body 3 to the crankcase 2 are screwed into the bolt-coupling portions 12 c, so that the bolt-coupling portions 12 c arranged in two front and rear locations with the cylinder bore axis A therebetween can equally bear a load produced by combustion pressure and the rigidity in coupling of the cylinder body 3 and the crankcase 2 can be improved.

Also, since at least one end portion of balancer shafts 22, 22′ arranged in the vicinity of and in parallel to the crankshaft 8 is born by the iron-alloy bearing members 12, 12′, the balancer shafts 22, 22′ can be enhanced in rigidity of bearing.

Furthermore, when the iron bearing brackets 12, 12′ are to be embedded in the crankcase 2 of aluminum alloy, upper-end surfaces 12 f of the bolt-coupling portions 12 c are not exposed to the cylinder-side mating surface 2 e of the crankcase 2 but arranged inwardly thereof, so that metallic members, which are different in hardness and material, are not intermixed into the coupling surfaces of the crankcase 2 and the cylinder block 3 and so degradation in sealing quality can be avoided. That is, if the upper-end surfaces 12 f of the iron bolt-coupling portions 12 c were brought into contact with a case-side mating surface 3 c, which is formed on the lower flange 3 b of the cylinder body 3 of aluminum alloy, degradation in sealing quality would be caused by a difference in coefficient of thermal expansion, or the like.

Also, since the bearing collar 12 having a larger outside diameter than that of the balancer drive gear 25 a is fitted onto an outer periphery of the bearing 11 a in the left bearing bracket 12, the balancer drive gear 25 a can be mounted without hindrance and without interfering a minimum inside-diameter portion of the boss portion 12 b of the bearing bracket 12 when the crankshaft 8 is mounted in the crankcase 2 in a state, in which the balancer drive gear 25 a is press-fitted (integral formation will do) onto and fixed to the crankshaft 8.

The crankcase 2 is of a two left and right split type to be composed of left and right case portions 2 a, 2 b. A left case cover 9 is detachably mounted on the left case portion 2 a, and a space surrounded by the left case portion 2 a and the left case cover 9 defines a flywheel-magneto chamber 9 a. A flywheel-magneto type generator 35 mounted on a left end of the crankshaft 8 is received in the flywheel-magneto chamber 9 a. In addition, the flywheel-magneto chamber 9 a is communicated to a chamber, in which the cam shaft is arranged, via chain chambers 3 d, 4 d described later, and a major part of a lubricating oil having lubricated a camshaft drops into the flywheel-magneto chamber 9 a via the chain chambers 3 d, 4 d.

Also, a right case cover 10 is detachably mounted on the right case portion 2 b, and a space surrounded by the right case portion 2 b and the right case cover 10 defines a clutch chamber 10 a.

Formed in front of the crankcase 2 is a crank chamber 2 c, and formed behind the crankcase is a transmission chamber 2 d. The crank chamber 2 c is opened toward the cylinder bore 3 a and substantially separated from other chambers such as the transmission chamber 2 d or the like. Therefore, pressure in the transmission chamber 2 d is varied due to up/down movements of the piston 6 to function as a pump.

The crankshaft 8 is arranged to receive left and right arm portions 8 a, 8 a and left and right weights 8 b, 8 b in the crank chamber 2 c. The crankshaft 8 is of an assembled type, in which a left crankshaft portion uniting the left arm portion 8 a, the left weight 8 b, and the left shaft portion 8 c is integrally coupled to a right crankshaft portion uniting the right arm portion 8 a, the right weight 8 b, and the right shaft portion 8 c′, through a cylindrical-shaped crank pin 8 d.

The left and right shaft portions 8 c, 8 c′ are rotatably supported through the crank bearings 11 a, 11 a′ on the left and right case portion 2 a, 2 b. The bearings 11 a, 11 a′ are press-fitted into the bearing holes 12 a of the iron-alloy bearing brackets 12, 12′, which are insert-cast into the left and right case portion 2 a, 2 b of aluminum alloy, as described above.

A transmission mechanism 13 is received and arranged in the transmission chamber 2 d. The transmission mechanism 13 comprises a main shaft 14 and a drive shaft 15, which are arranged in parallel to the crankshaft 8, and is of a constantly meshing type construction, in which first to fifth speed gears 1 p to 5 p mounted on the main shaft 14 constantly mesh with first to fifth speed gears 1 w to 5 w mounted on the drive shaft 15.

The main shaft 14 is supported through main-shaft bearings 11 b, 11 b by the left and right case portions 2 a, 2 b, and the drive shaft 15 is supported through drive-shaft bearings 11 c, 11 c by the left and right case portions 2 a, 2 b.

A right end portion of the main shaft 14 extends through the right case portion 2 b to project rightward, and the clutch mechanism 16 described above is mounted on the projecting portion, the clutch mechanism 16 being positioned in the clutch chamber 10 a. A large reduction gear (input gear) 16 a of the clutch mechanism 16 meshes with a small reduction gear 17, which is fixed to and mounted on a right end portion of the crankshaft 8.

A left end portion of the drive shaft 15 projects outward from the left case portion 2 a, and a drive sprocket 18 is mounted on the projecting portion. The drive sprocket 18 is coupled to a driven sprocket of a rear wheel via a drive chain.

The balancer device 19 according to the embodiment comprises front and rear (first and second) balancer devices 20, 20′, which are arranged in a manner to interpose therebetween the crankshaft 8 and are of substantially the same construction. The front and rear balancer devices 20, 20′ comprise front and rear (first and second) balancer shafts 22, 22′ of non-rotation type, and front and rear (first and second) balancer weights 24, 24′ rotatably supported through bearings 23, 23 on the balancer shafts.

Here, the front and rear balancer shafts 22, 22′ also serve as case bolts (coupling bolts) for clamping and coupling the left and right case portion 2 a, 2 b together in a direction along the crankshaft. The front and rear balancer shafts 22, 22′ comprise a case interior portion 22 g positioned to extend through the crank chamber 2 c, and a bearing portion extending outside of the crank chamber 2 c to bear the balancer weight 24, 24′ through the bearing 23. While the bearing portion is set to have a relatively large diameter, which involves no hindrance in bearing the balancer weight 24, 24′ through the bearing 23, the case interior portion 22 g is set to have a small diameter in a manner to withstand a required coupling force, that is, substantially the same as that of ordinary case coupling bolts. In addition, a boundary portion between the case interior portion 22 g of a small diameter and the bearing portion of a large diameter is formed to be tapered.

Flanges 22 a, 22 a′ formed inwardly of the supported balancer weights 24, 24′ in a widthwise direction of the engine abut against boss portions 12 g of the bearing brackets 12, 12′ inserted into the left and right case portion 2 a, 2 b and fixing nuts 21 b, 21 a are screwed onto opposite end portions of the balancer shafts whereby the balancer shafts 22, 221 couple the left and right case portion 2 a, 2 b together.

The balancer weights 24, 24′ of the front and rear balancer devices 20, 20′ comprise a substantially semi-circular weight body 24 a, 24 a′ and a circular-shaped gear support 24 b, 24 b′ formed integral therewith, and a ring-shaped balancer driven gear 24 c, 24 c′ is mounted on and fixed to the gear support 24 b, 24 b′. In addition, the reference numeral 24 d denotes a lightening hole, by which a weight of a portion diametrically opposed to the weight body 24 a, 24 a′ is reduced as much as possible.

Here, the front and rear balancer devices 20, 20′ are arranged to approach the crankshaft 8 to such an extent that when viewed in a direction along the crankshaft, loci of rotation of the balancer weights 24, 24′ partially overlap a locus of rotation of a conrod coupling portion formed by coupling a large end 7 a of the conrod 7 to the arm portion 8 a of the crankshaft 8 by a crank pin 65, and a locus of rotation of the right weight 8 b.

The balancer driven gear 24 c′ of the rear balancer device 20′ meshes with the balancer drive gear 25 a on the crankshaft 8. The balancer drive gear 25 a is mounted on the gear body 25, which is press-fitted onto and fixed to the left shaft portion 8 c of the crankshaft 8 to be rotatable relative thereto. In addition, the gear body 25 is adjacent to the left journal bearing 11 a.

In addition, the reference numeral 25 b denotes a sprocket integrally formed on the gear body 25 to drive a timing chain, the sprocket 25 b having on an outer side thereof a registration mark 25 c for setting of a valve timing as shown in FIG. 11. The gear body 25 is fitted onto the crankshaft 8 such that when the crankshaft 8 is in a top dead center position of compression, the registration mark 25 c coincides with the cylinder bore axis A as viewed in the direction along the crankshaft.

Also, the balancer driven gear 24 c mounted on the gear support 24 b of the balancer weight 24 of the front balancer device 20 meshes with a balancer drive gear 17 a on the crankshaft 8. The balancer drive gear 17 a is supported on the small reduction gear 17, which is fixed to and mounted on the right shaft portion 8 c′ of the crankshaft 8, to be rotatable relative thereto.

Here, the balancer drive gear 25 a for the rear balancer device 20′ is supported on the gear body 25 to be rotatable relative thereto, and the balancer drive gear 17 a for the front balancer device 20 is supported on the small reduction gear 17 to be rotatable relative thereto. And buffer springs 33 formed from U-shaped plate springs are interposed between the balancer drive gears 25 a, 17 a and the gear body 25 and the small reduction gear 17, whereby shocks caused by torque nonuniformity of the engine or the like are suppressed in transmission to the front and rear balancer devices 20, 20′.

Here, the balancer drive gear 17 a for the front-side driving will be described in further detail with reference to FIG. 14, and the same is with the balancer drive gear 25 a for the rear-side driving. The balancer drive gear 17 a is ring-shaped and supported on a slide surface 17 b, which is formed on a side of the small reduction gear 17 to have a smaller diameter than that of the same, to be rotatable relative thereto. A multiplicity of U-shaped spring holding grooves 17 c are recessed in the slide surface 17 b to extend radially about a core of the crankshaft, the U-shaped buffer springs 33 being inserted and arranged in the spring holding grooves 17 c. Open ends 33 a, 33 a of the buffer springs 33 are latched on front and rear steps of latch recesses 17 d formed on an inner peripheral surface of the balancer drive gear 17 a.

When relative rotations due to torque nonuniformity or the like generate between the small reduction gear 17 and the balancer drive gear 17 a, the buffer springs 33 are elastically deformed in a direction, in which spacings between the ends 33 a, 33 a decrease, to absorb torque nonuniformity. In addition, the reference numeral 17 g denotes a cover plate for holding the buffer springs 33 in the holding grooves 17 c, 17 h a key for coupling the small reduction gear 17 to the crankshaft 8, and 17 e, 17 f registration marks used at the time of assembly of the small reduction gear 17 and the balancer drive gear 17 a.

Mechanisms are provided on the front and rear balancer devices 20, 20′ to regulate backlash between the balancer driven gears 24 c, 24 c′ and the balancer drive gears 17 a, 25 a. The regulating mechanisms are constructed to make balancer axes of the balancer shafts 22, 22′ and axes of rotation of the balancer driven gears 24 c, 24 c′ slightly eccentric relative to each other. More specifically, when the balancer shafts 22, 22′ are turned about the balancer axes, the eccentricity causes slight changes in spacings between axes of rotation of the balancer driven gears 24 c, 24 c′ and axes of rotation of the balancer drive gears 17 a, 25 a, so that changes in backlash result.

Here, the mechanisms for turning the front and rear balancer shafts 22, 22′ about the balancer axes are different between the front balancer device 20 and the rear balancer device 20′. First, with the rear balancer device 20′, a hexagonal-shaped latch projection 22 b is formed on a left end of the rear balancer shaft 22′, and a spline-shaped (polygonal star-shaped) latch hole 26 a formed on one end of a turning lever 26 is latched on the latch projection 22 b. Also, a bolt hole 26 b being arcuate about the balancer axis is formed on the other end of the turning lever 26.

A fixing bolt 27 inserted into the bolt hole 26 b is implanted into a guide plate 28. The guide plate 28 is generally arcuate to be bolted and fixed to the crankcase 2. In addition, the guide plate 28 also has a function of regulating a flow of a lubricating oil.

Backlash of the rear balancer device 20′ is regulated by turning the turning lever 26 in a state, in which the fixing nut 21 a is loosened, to put backlash in an appropriate condition, and then using the fixing bolt 27 a and the nut 27 b to fix the turning lever 26, and thereafter the fixing nut 21 a is clamped.

Formed on a left end of the front balancer shaft 22 is a grip portion 22 f having an oval-shaped cross-section and formed on both sides of a circular cross section with flat portions 22 e (see FIG. 12). A collar 29 a having an inner peripheral shape corresponding to an outer peripheral shape of the grip portion 22 f is mounted on the grip portion 22 f, and a holding portion 29 b of a holding lever 29 is mounted outside the collar 29 a to be axially movable but non-rotatable relative thereto. A tip end 29 e of the holding lever 29 is fixed to a boss portion 2 f of the left case portion 2 a by means of a bolt 29 f. Also, a clamping slit 29 c is formed in the holding portion 29 b of the holding lever 29 to prevent rotation of the collar 29 a and hence the balancer shaft 22 upon clamping of a fixing bolt 29 d. Moreover, the fixing nut 21 b is screwed axially outwardly of the collar 29 a onto the balancer shaft 22 with a washer therebetween.

Backlash of the front balancer device 20 is regulated by loosening, preferably, removing the fixing nut 21 b, using a tool to grip the grip portion 22 f of the balancer shaft 22 to turn the same in a manner to put backlash in an appropriate condition, and then clamping the fixing bolt 29 d, and thereafter the fixing nut 21 b is clamped.

Also, an upper portion of the latch projection 22 b on the balancer shaft 22, 22′ is arcuately cut out to form a lubricating oil introducing portion 22 c. A guide hole 22 d is opened to the introducing portion 22 c, the guide hole 22 d being extended through the balancer shaft 22 to reach a lower portion of an outer peripheral surface of the balancer shaft, whereby the lubricating oil introducing portion 22 c is communicated to an inner peripheral surface of the balancer bearing 23. In this manner, a lubricating oil dropping into the lubricating oil introducing portion 22 c is fed to the balancer bearing 23.

Here, the balancer weight 24 and the balancer driven gear 24 c are arranged in a right end in the direction along the crankshaft with the front balancer device 20 while they are arranged in a left end with the rear balancer device 20′. Also, the balancer driven gear 24 c, 24 c′ is positioned rightwardly of the balancer weight 24, 24′ with both the front and rear balancer devices 20, 20′, so that the front and rear balancer weights 24, 24′ and the balancer driven gears 24 c, 24 c′ are set to assume the same shapes with both the front and rear balancer devices.

In this manner, according to the embodiment, since the weight body 24 a and the balancer driven gear 24 c of the balancer weight 24 are arranged in the right end (one end) of the front (first) balancer shaft 22 in the direction along the crankshaft and the weight body 24 a′ and the balancer driven gear 24 c′ of the balancer weight 24′ are arranged in the left end (the other end) of the rear (second) balancer shaft 22′ in the direction along the crankshaft, heavy loads are arranged left and right across the direction along the crankshaft whereby it is possible to avoid that worsening in weight balance in the direction along the crankshaft, which is caused in the case where a two-axis type balancer device is provided.

Also, since the front and rear balancer shafts 22, 22′ are arranged to approach the crankshaft 8 to such an extent that loci of rotation of the balancer weights 24, 24′ partially overlap a locus of rotation of that portion of the crankshaft 8, which is coupled to the conrod, it is possible to avoid that large-sizing around the crankcase, which is caused by arrangement of two balancer shafts away from a crankshaft, in the case where a two-axis type balancer device is provided.

Also, since the front and rear balancer shafts 22, 22′ also serve as case bolts for coupling the left and right case portions 2 a, 2 b together, it is possible to enhance a crankcase in coupling rigidity while suppressing complication of a construction and an increase in the number of parts, in the case where a two-axis type balancer device is adopted, and to reduce a space, in which exclusive case coupling bolts are arranged, and in this point it is possible to avoid large-sizing around the crankcase.

Moreover, since the crankcase-interior portion 22 g of the balancer shaft 22, 22′ is formed to have a smaller diameter than the born portion, there is caused no hindrance in bearing the balancer weight 24, 24′, and therefore it is possible to make the balancer shaft 22, 22′ close to the crankshaft 8 to the limit, and in this point it is also possible to avoid large-sizing around the crankcase.

Also, since the weight bodies 24 a, 24 a′, respectively, are united with the balancer gears 24 c, 24 c′, respectively, and rotatably supported on the balancer shafts 22, 22′, it suffices to rotatingly drive only the balancer weights 24, 24′ composed of the weight body 24 a and the balancer driven gear 24 c, and it is unnecessary to rotatingly drive the balancer shafts themselves, according to which it is possible to achieve making effective use of an engine output.

Moreover, since the front and rear balancer weights 24, 24′ are arranged to be close to outsides of the journal bearings 11 a, 11 a′, which support both sides of the arm portion 8 a constituting that portion of the crankshaft 8, which is coupled to the conrod, a dimension in the direction along the crankshaft can be minimized, so that in the case where a two-axis type balancer device is provided, it is possible to avoid a problem that enlargement around the crankcase in the direction along the crankshaft is caused due to arrangement of balancer weights away from that portion of a crankshaft, which is coupled to a conrod, in a direction along the crankshaft.

Also, since the balancer weights 24, 24′ are separate from the balancer shafts 22, 22′, freedom in assembly can be improved as compared with a constitution, in which these elements are united together.

Also, since axes of rotation of the balancer driven gears 24 c, 24 c′ are made eccentric from axes of the balancer shafts 22, 22′, a simple construction and a simple operation of rotating the balancer shafts make it possible to regulate backlash relating to the balancer driven gears 24 c, 24 c′ and the balancer drive gears 17 a, 25 a on the crankshaft 8, thus enabling preventing generation of noise.

With the front balancer shaft 22, backlash is regulated by using a tool to grip the grip portion 22 f formed leftward in a lateral direction of the balancer shaft 22 to turn the balancer shaft 22, and with the rear balancer shaft 22′, backlash is regulated by turning the turning lever 26 provided similarly leftward in the lateral direction. In this manner, with either of the front and rear balancer shafts 22, 22′, backlash can be regulated from leftwardly of the engine, so that it is possible to efficiently perform a work of backlash regulation.

Also, since the balancer drive gear 17 a on the crankshaft 8 to mesh with the balancer driven gear 24 c is arranged to be rotatable relative to the slide surface 17 b of the small reduction gear 17 fixed to the crankshaft 8 and the U-shaped buffer springs 33 are arranged in the spring holding grooves 17 c recessed in the slide surface 17 d, such compact construction absorbs shocks caused by torque nonuniformity of the engine or the like to be able to smoothly actuate the balancer device. In addition, the same is with the balancer drive gear 25 a.

Moreover, a cooling water pump 48 is arranged on a right end of the front balancer shaft 22 to be coaxial therewith. An Oldham's coupling or the like, having the same structure as that with a lubricating oil pump 52 described later connects a rotating shaft of the cooling water pump 38 to the balancer shaft 22 to accommodate some off-centering between the both.

A valve gear in the embodiment causes the crankshaft 8 to drivingly rotate an intake cam shaft 36 and an exhaust cam shaft 37 arranged in the head cover 5. Concretely, the crank sprocket 25 b on the gear body 25 press-fitted onto the left shaft portion 8 c of the crankshaft 8 and an intermediate sprocket 38 a born by a support shaft 39 implanted in the cylinder head 4 are coupled to each other by a timing chain 40, and an intermediate gear 38 b formed integral with the intermediate sprocket 38 a and having a smaller diameter than that of the intermediate sprocket 38 a meshes with intake and exhaust gears 41, 42, which are fixed to ends of the intake and exhaust cam shafts 36, 37. In addition, the timing chain 40 is disposed to pass through chain chambers 3 d, 4 d formed on left walls of the cylinder block 3 and the cylinder head 4.

The intermediate sprocket 38 a and the intermediate gear 38 b are born through two sets of needle bearings 44 by the support shaft 39, which is disposed on the cylinder bore axis A to extend through the chain chamber 4 d of the cylinder head 4 in the direction along the crankshaft. A flange 39 a of the support shaft 39 is fixed to the cylinder head 4 by means of two bolts 39 b. In addition, the reference numerals 39 c, 39 d denote gaskets for sealing.

Here, commercially available products (standardized goods) are employed for the two sets of needle bearings 44, 44, a collar 44 a for regulation of spacing is arranged between the respective bearings 41, 41, and thrust washers 44 b, 44 b are arranged on both ends to bear thrust loads. The thrust washers 44 b are stepped to comprise a large-diameter portion, which slidingly contacts with an end surface of the intermediate sprocket, and a stepped portion projecting axially toward the needle bearing 44.

In this manner, since the collar 44 a for regulation of spacing is interposed between the two sets of needle bearings 44, 44, regulation of a length of the collar 44 a makes it possible to employ commercially available standardized goods for the needle bearings, thus enabling reduction in cost.

Also, since stepped ones are employed for the thrust washers 44 b, it is possible to improve workability in assembling the intermediate sprocket 38 a. More specifically, when the intermediate sprocket 38 a is to be assembled, the support shaft 39 is inserted from outside in a state, in which the intermediate sprocket 38 a and the intermediate gear 38 b with the thrust washers disposed on both ends thereof in a manner to prevent falling of the thrust washers are arranged in the chain chamber 4 d. By latching stepped portions of the thrust washers 44 b on a shaft hole of the intermediate sprocket 38 a, falling of the thrust washers 44 b can be prevented, and so the assembling quality can be improved correspondingly.

Also, formed in the support shaft 39 are oil holes 39 e, through which a lubricating oil introduced from the cam chamber via an oil introduction hole 4 e formed in the cylinder head 4 is fed to the needle bearings 44.

Also, four lightening holes 38 c and two lightening/peeping holes 38 c′ for assembly are formed on the intermediate sprocket 38 a to be spaced 60 degrees from one another. And registration marks 38 d are marked on teeth disposed substantially centrally of the peeping holes 38 c′ of the intermediate gear 38 b, and registration marks 41 a, 42 a are marked on two teeth of the intake and exhaust cam gears 41, 42 to correspond to the registration marks 38 d. Here, when the left and right registration marks 38 d, 38 d are made coincident with the registration marks 41 a, 42 a, the intake and exhaust cam shafts 41, 42 are positioned in the top dead center position of compression.

Moreover, registration marks 38 e, 38 e are formed on those portions of the intermediate sprocket 38 a, which are positioned on a cover-mating surface 4 f of the cylinder head 4 when the registration marks 38 d are coincident with the registration marks 41 a, 42 a.

In order to set a valve timing, the registration mark 25 c (see FIG. 11) is first made to coincide with the cylinder bore axis A, whereby the crankshaft 8 is kept in the top dead center position of compression. Also, the intermediate sprocket 38 a and the intermediate gear 38 b, which are mounted on the cylinder head 4 with the support shaft 39 therebetween, are positioned so that the registration marks 38 e, 38 e on the intermediate sprocket 38 a are made coincident with the cover-mating surface 4 f, and in this state, the cam sprocket 25 b and the intermediate sprocket 38 a are coupled to each other by the timing chain 40. And the intake and exhaust cam gears 41, 42 on the intake and exhaust cam shafts 36, 37 are caused to mesh with the intermediate gear 38 b while it is confirmed through the peeping holes 38 c′ whether the registration marks 41 a, 42 a are coincident with the registration marks 38 d of the intermediate gear 38 b, and the intake and exhaust cam shafts 36, 37 are fixed to an upper surface of the cylinder head 4 through a cam carrier.

In this manner, since the lightening/peeping holes 38 c′ are provided on the intermediate sprocket 38 a having a large diameter to enable confirming therethrough a state, in which the registration marks 38 d of the intermediate gear 38 b disposed on a back surface side and having a small diameter are coincident with the registration marks 41 a, 42 a of the cam gears 41, 42, a position, in which the intermediate gear 38 b meshes with the cam gears 41, 42, can be visually confirmed with ease and certainty and the valve timing can be set without hindrance although the intermediate gear 38 b having a small diameter is arranged on a back surface of the intermediate sprocket 38 a having a large diameter.

Also, since the intermediate gear 38 b can be arranged on a back surface side of the intermediate sprocket 38 a, it is possible to shorten dimensions from the cam gears 41, 42, which mesh with the intermediate gear 38 b, to cam noses 36 a, and to correspondingly decrease twist angles of the cam shafts, thus enabling enhancing an accuracy, with which valve opening and closing timings are controlled. Also, parts around the cam shafts can be made compact.

More specifically, for example, in the case where the intermediate gear 38 b is arranged on a front surface of the intermediate sprocket 38 a, a valve timing can be readily set but the dimensions from the cam gears 41, 42 to the cam noses are lengthened, and the twist angles of the cam shafts are correspondingly increased, so that the accuracy, with which valve opening and closing timings are controlled, is degraded.

Also, in the case where the intermediate gear 38 b is arranged on the front surface of the intermediate sprocket 38 a, it is necessary to enlarge spacings between the intermediate sprocket support shaft 39 and the cam shafts 36, 37 in order to avoid interference between the intermediate sprocket 38 a and the cam shafts 36, 37, and there is correspondingly caused a fear of parts around the cam shafts being made large-sized.

Here, a backlash regulation mechanism is provided between the intermediate gear 38 b and the cam gears 41, 42. The regulation mechanism is constructed such that an intake cam gear 41 and an exhaust cam gear 42, respectively, comprise two gears, that is, a drive gear (power electrically-driven gear) 46, and a shift gear (regulation gear) 45, and the drive gear 46 and the shift gear 45 are adjustable in angular position.

More specifically, the shift gear 45 and the drive gear 46 are fixed to flanges 36 b, 37 b formed on ends of the cam shafts 36, 37 to be made adjustable in angular position by means of four circumferentially-lengthy slots 45 a, 46 a and four lengthy bolts 68 a, and relief portions 46 b are cut out to be formed on the outwardly-positioned drive gears 46, so that with the use of the relief portions 46 b only the shift gears 45 are fixed in a manner to be made adjustable in angular position by means of two lengthy slots 45 b and two short bolts 68 b.

Backlash is regulated in the following procedure. In addition, with the engine according to the embodiment, the intermediate gear 38 b rotates in a counterclockwise direction as viewed from leftwardly of the engine as shown in FIG. 3. Accordingly, both the intake and exhaust cam gears 41, 42 rotate in a clockwise direction. While backlash regulation is described herein with respect to the intake cam gear 41, the same is with the exhaust cam gear 42.

First, all of fixing bolts 68 a, 68 b on the intake cam gear 41 are loosened, the shift gear 45 is turned clockwise, a clockwise leading tooth-surface of the shift gear 45 is caused to lightly abut against a counterclockwise trailing tooth-surface of the intermediate gear 38 b, and in this state two short bolts 68 b are used to fix the shift gear 45 to the flange 36 b of the cam shaft 36. Then the drive gear 46 is turned counterclockwise, a counterclockwise leading tooth-surface (driven surface) thereof is caused to abut against a counterclockwise leading tooth-surface (drive surface) of the intermediate gear 38 b so that a predetermined backlash is obtained, and in this state four lengthy bolts 68 a are driven to thereby fix the drive gear 46 and the shift gear 45 to the intake cam shaft 36.

In this manner, since the intake cam gear 41 and the exhaust cam gear 42, respectively, comprise the drive gear (power transmitting gear) 46 and the shift gear (regulation gear) 45 rotatable relative to the drive gear, backlash can be regulated by turning the shift gear 45 relative to the drive gear 46 forward or rearward in a direction of rotation.

In addition, while the embodiment has been described with respect to the case where both the drive gear 46 and the shift gear 45, which constitute the cam gears 41, 42, are rotatable relative to the cam shafts, one of the drive gear 46 and the shift gear 45 may be made relatively rotatable and the other gear may be made integral with the cam shaft. In this case, it is desired that the gears made integral with the cam shafts be made power transmitting gears. Even with such constitution, the same action and effect as those in the embodiment are obtained.

Also, while the embodiment has been described with respect to the case where the invention is applied to a chain-driving system, the invention can be of course applied to a drive-system with a toothed belt.

Subsequently, a lubrication construction will be described.

With a lubricating device 50 in the engine according to the embodiment, a lubricating oil stored in a separate lubricating oil tank 51 is drawn and pressurized by a lubricating oil pump 52 through down tubes 56 c of a car body frame 56, an oil discharged from the pump 52 is divided into three systems of a cam lubricating system 53, a transmission lubricating system 54, and a crank lubricating system 55 to be fed to respective sections being lubricated, and the lubricating oil having lubricated these sections being lubricated is returned to the lubricating oil tank 51 making use of pressure fluctuation in the crank chamber 2 c, caused by up and down movements of the piston 6.

The lubricating oil tank 51 is formed integrally in a space surrounded by a head pipe 56 a of the car body frame 56, a main tube 56 b, the down tubes 56 c, and a reinforcement bracket 56 d. The lubricating oil tank 51 is communicated to a cross pipe 56 e, which connects lower portions of the down tubes 56 c together, from the down tubes 56 c.

The cross pipe 56 e is connected to a suction port of the lubricating oil pump 52 through an outlet tube 56 f, which is connected to the cross pipe, an oil hose 57 a, a joint pipe 57 b, and an oil suction passage 58 a formed on the crankcase cover. A discharge port of the lubricating oil pump 52 is connected to an oil filter 59 via an oil discharge passage 58 b, an outside connection chamber 58 c, and an oil passage 58 d, and divided on a downstream side of the oil filter 59 into the three lubrication systems 53, 54, 55.

The oil filter 59 comprises an oil element 59 e arranged in a filter chamber 59 d, which comprises a filter cover 47 detachably mounted on a filter recess 10 b recessed in the right case cover 10.

The cam lubricating system 53 has an outline construction such that a lower end of a longitudinal portion 53 a of a T-shaped lubricating oil pipe is connected to a cam-side outlet 59 a of an oil passage formed outside the filter recess 10 b from the filter cover 47 and left and right ends of a transverse portion 53 b of the lubricating oil pipe are connected to cam-shaft oil-feed passages 53 c to permit a lubricating oil to be fed to sections being lubricated, such as bearing portions of the cam shafts 36, 37, via the passages 53 c.

The transmission lubricating system 54 is constructed in the following manner. A right transmission oil-feed passage 54 a formed in the right case portion 2 b is connected to a transmission-side outlet 59 b of the oil filter 59, and the oil-feed passage 54 a is communicated to a main-shaft hole 14 a formed in an axial core of the main shaft 14 via a left transmission oil-feed passage 54 b formed in the left case portion 2 a. And the main-shaft hole 14 a is communicated to sliding portions of the main shaft 14 and change gears via a plurality of branch holes 14 b, so that a lubricating oil fed to the main-shaft hole 14 a is fed to the sliding portions through the branch holes 14 b.

Also, an intermediate portion of the left transmission oil-feed passage 54 b is communicated to a bolt hole 60 a, into which a case bolt 60 coupling the left and right case portions 2 a, 2 b to each other is inserted. The bolt hole 60 a comprises a hole having a somewhat larger inside diameter than an outside diameter of the case bolt 60 and formed in cylindrical-shaped boss portions 60 c, 60 c, which are formed on the left and right case portions 2 a, 2 b in a manner to face and abut against mating surfaces of the case portions. The boss portions 60 c are positioned in the vicinity of meshing portions of gear trains on the main shaft 14 and the drive shaft 15, and formed with a plurality of branch holes 60 b, through which a lubricating oil in the bolt hole 60 a is jetted toward the meshing portions. In addition, while the bolt 60 shown in FIG. 19 is described with the left and right case portions developed, these indicate the same bolt.

Moreover, a right end of the bolt hole 60 a is communicated to a drive-shaft hole 15 a formed in an axial core of the drive shaft 15 via a communication hole 54 c. The drive-shaft hole 15 a is closed at its left portion by a partition wall 15 c, and communicated to sliding portions of the drive shaft 15 and drive gears through a plurality of branch holes 15 b. In this manner, a lubricating oil fed to the drive-shaft hole 15 a is fed to the sliding portions through the branch holes 15 b.

The crank lubricating system 55 is constructed in the following manner. Formed in the filter cover 47 is a crank oil-feed passage 55 a extending toward the lubricating oil pump 52 from a crank-side outlet 59 c, the passage 55 a being communicated to a communication hole 62 a formed through an axial core of a rotating shaft 62 of the lubricating oil pump 52, and the communication hole 62 a is communicated via a connection pipe 64 to a crank oil-feed hole 8 e formed in an axial core of the crankshaft 8. And the crank oil-feed hole 8 e is communicated to a pin hole 65 a of the crank pin 65 via a branch hole 8 f, the pin hole 65 a being opened through the branch hole 65 b to rolling surfaces of a needle bearing 7 b on the large end 7 a of the conrod 7. In this manner, a lubricating oil having been filtered through the oil filter 59 is fed to the rolling surfaces of the needle bearing 7 b.

The lubricating oil pump 52 is constructed in the following manner. A pump chamber 61 c is recessed in a right casing 61 b of a split-type casing 61 comprising left and right casings 61 a, 61 b, and a rotor 63 is rotatably arranged in the pump chamber 61 c. The oil pump has an outline construction such that the rotating shaft 62 is inserted and arranged in an axial core of the rotor 63 to extend axially thereof, and the rotating shaft 62 and the rotor 63 are fixed to each other by means of a pin 63 a. In addition, the oil suction passage 58 a and the oil discharge passage 58 b, respectively, are connected to upstream and downstream sides of the pump chamber in the left casing 61 a. In addition, the reference numeral 66 denotes a relief valve for keeping a discharge pressure of the lubricating oil pump 52 to a predetermined value or lower, the relief valve permitting pressure on a discharge side of the lubricating oil pump 52 to escape toward the oil suction passage 58 a when the pressure assumes the predetermined value or higher.

The rotating shaft 62 is cylindrical in shape to extend axially through the pump casing 61, and a right end of the shaft shown in the figure is opened to the crank oil-feed passage 55 a. Also, a power transmission flange 62 b is formed integral with a left end of the rotating shaft 62 shown in the figure. The flange 62 b is opposed to a right end surface of the crankshaft 8, and the flange 62 b and the crankshaft 8 are connected to each other such that an Oldham's coupling 67 can accommodate some off-centering therebetween.

More specifically, the Oldham's coupling 67 is constructed such that a coupling plate 67 a is arranged between the crankshaft 8 and the flange 62 b and a pin 67 b implanted into an end surface of the crankshaft 8 and a pin 67 c implanted into the flange 62 b are inserted into connection holes 67 d of the coupling plate 67 a.

Also, the connection pipe 64 serves to communicate a right-end opening of the crankshaft 8 and a left-end opening of the rotating shaft 62 to each other, and oil seals 64 a provide sealing between inner peripheries of the opening of the crankshaft and the opening of the rotating shaft and an outer periphery of the connection pipe 64.

Here, as described above, the crank chamber 2 c is defined with the other transmission chamber 2 d, flywheel-magneto chamber 9 a, and the clutch chamber 10 a, whereby an oil returning mechanism is constituted, in which a stroke of the piston 6 causes pressure in the crank chamber 2 c to fluctuate between positive and negative and such pressure fluctuation causes the lubricating oil in the respective chambers to be returned to the lubricating oil tank 51.

More specifically, a discharge port 2 g and a suction port 2 h are formed on the crank chamber 2 c. Arranged at the discharge port 2 g is a discharge-port reed valve 69, which opens when pressure in the crank chamber is positive, and arranged at the suction port 2 h is a suction-port reed valve 70, which opens when pressure in the crank chamber is negative.

And the discharge port 2 g is communicated to the clutch chamber 10 a via a communication hole 2 i from the crank chamber 2 c to be communicated to the transmission chamber 2 d via a communication hole 2 j from the clutch chamber 10 a, and the transmission chamber 2 d is further communicated to the flywheel-magneto chamber 9 a via a communication hole 2 k. Are turn port 2 m formed in communication with the flywheel-magneto chamber 9 a is communicated to the lubricating oil tank 51 via a length of return hose 57 c, an oil strainer 57 d, and a length of return hose 57 e.

Here, a guide plate 2 n is arranged at the return port 2 m. The guide plate 2 n decreases a gap a of the return port 2 m between the guide plate and a bottom wall 2 p and ensures a wide width b to thereby function to surely discharge the lubricating oil.

Also, connected to the lubricating oil tank 51 is an oil separation mechanism for separating an oil mist contained in an air in the tank with a centrifugal force to return the same to the crank chamber 2 c. The oil separation mechanism is constructed such that an introduction hose 72 a connected at one end thereof to a top of the lubricating oil tank 51 is connected at the other end thereof tangentially to an upper portion of a cone-shaped separation chamber 71 and a return hose 72 b connected to a bottom of the separation chamber 71 is connected to the suction port 2 h of the crank chamber 2 c. In addition, the air, from which the oil mist has been separated, is discharged to the atmosphere via an exhaust hose 72 c.

As described above, according to the embodiment, since the crank chamber 2 c is made a substantially closed space so as to make pressure therein fluctuate upon up and down movements of the piston 6, and the lubricating oil having flown into the crank chamber 2 c is fed to the lubricating oil tank 51 according to fluctuation of pressure in the crank chamber 2 c it is possible to dispense with any exclusive oil-feed pump (scavenge pump), thus enabling achieving simplification in construction and reduction in cost.

Also, since the discharge-port reed valve (outflow-side check valve) 69 is arranged in the vicinity of a connection of the oil-feed passage of the crank chamber 2 c to open when pressure in the crank chamber is increased and to close when pressure in the crank chamber is decreased, the lubricating oil in the crank chamber can be more surely fed to the lubricating oil tank 51.

Also, since a portion above an oil level in the lubricating oil tank 51 is connected to the crank chamber 2 c via the return hoses 72 a, 72 b and the discharge-port reed valve (suction-side check valve) 70 is arranged in the vicinity of a connection of the return hose of the crank chamber 2 c to open when pressure in the crank chamber 2 c is decreased and to close when pressure in the crank chamber is increased, a necessary air is drawn into the crank chamber 2 c when the piston ascends and pressure in the crank chamber 2 c is increased when the piston descends, so that the lubricating oil in the crank chamber 2 c can be more surely fed.

Incidentally, in the case where there is provided no path, through which an air from outside is supplied into the crank chamber, there is a fear that when sealing between a piston and a cylinder bore is high in level, pressure in a crank chamber becomes negative upon upward movement of a piston and becomes negative or positive in low level even when the piston descends, whereby oil feed is made impossible.

Moreover, since the centrifugal-type lubricating-oil mist separator 71 for separation of a lubricating-oil mist is provided midway the return passages 72 a, 72 b, the separated lubricating oil is returned to the crank chamber 2 c via the return hose 72 b, and an air, from which the mist content is separated, is discharged to the atmosphere, only the mist content of the lubricating oil can be returned into the crank chamber, so that it is possible to avoid that decrease in oil-feed efficiency, which is caused by flowing of excess air into the crank chamber, and to surely feed a lubricating oil in the crank chamber while preventing atmospheric pollution.

Also, since the lubricating oil pump 52 is arranged to be connected to one end of the crankshaft 8 and a discharge port of the lubricating oil pump 52 is communicated to the crank oil-feed hole (oil-feed passage in the crankshaft) 8 e formed in the crankshaft 8 via the communication hole (oil-feed passage in the pump) 62 a formed in the lubricating oil pump 52 and the connection pipe 64, a simple and compact construction makes it possible to feed a lubricating oil to sections being lubricated, of the crankshaft 8.

Also, since the crankshaft 8 and the lubricating oil pump 52 are connected to each other by the Oldham's coupling 67 capable of accommodating displacement in a direction perpendicular to the crankshaft, the communication hole 62 a and the crank oil-feed hole 8 e are connected to each other by the connection pipe 64, and the O-rings 64 a having elasticity are interposed between the connection pipe 64 and the communication hole 62 a and the crank oil-feed hole 8 e, the lubricating oil can be fed to the sections being lubricated, without hindrance even when some off-centering is generated between the crankshaft 8 and the pump shaft 62, so that a necessary lubricating quality can be ensured.

Moreover, since the cylindrical-shaped boss portions 60 c are formed in the vicinity of the main shaft 14 and the drive shaft 15, which constitute a transmission, the case bolt 60 for crankcase coupling is inserted into the bolt hole 60 a thereof to have a space between the internal surface of the bolt hole 60 a and the external surface of the case bolt 60 serving as a lubricating oil passage, and the branch holes (lubricating oil feed holes) 60 b are formed in the boss portions 60 c to be directed toward the transmission gears on the boss portions 60 c, a lubricating oil can be fed to meshing surfaces of the transmission gears without the provision of exclusive lubricating oil feed passages.

Also, since the other end of the lubricating oil passage defined between the internal surface of the bolt hole 60 c and the external surface of the case bolt 60 is communicated to the counter-output side opening of the drive-shaft hole (lubricating oil passage) 15 a formed in the drive shaft 15, a lubricating oil can be fed to sliding portions of the transmission gears on the drive shaft 15 without the provision of exclusive lubricating oil feed passages.

FIG. 22 is a view illustrating an embodiment of the invention of claim 7, in which a lubricating oil can be more surely fed to the balancer bearing 23. In the figure, the same reference numerals as those in FIG. 5 denote the same or corresponding elements.

In the embodiment, formed in balancer bearings 22, 22′ is an oil introduction passage 80 to communicate the crank chamber 2 c and mounts of the bearings 23 for the balancer weights 24, 24′ to each other. An oil inlet 80 a of the oil introduction passage 80 is opened to the crank chamber 2 c, and an oil outlet 80 b is opened toward the bearing 23.

Pulsation of pressure in the crank chamber 2 c caused by reciprocation of the piston 6 causes a lubricating oil to be introduced into the oil introduction passage 80 from the oil inlet 80 a and to be discharged from the oil outlet 80 a. Thereby, the lubricating oil is surely fed to the bearings 23 to improve the lubricating quality of the bearings 23.

INDUSTRIAL APPLICABILITY

According to the invention of claim 1, since the first balancer weight is arranged on one end of the first balancer shaft in a direction along the crankshaft and the second balancer weight is arranged on the other end of the second balancer shaft in the direction along the crankshaft, two heavy balancer weights are arranged on both sides across the direction along the crankshaft to be made left-right symmetric whereby it is possible to avoid that worsening in weight balance in the direction along the crankshaft, which is caused in the case where a two-axis type balancer device is provided.

Also, since the first and second balancer shafts are made close to the crankshaft so that loci of rotation of the first and second balancer weights partially overlap a locus of rotation of that portion of the crankshaft, which is coupled to a conrod when viewed in the direction along the crankshaft, it is possible to avoid a problem that large-sizing around the crankcase is caused due to arrangement of balancer shafts away from a crankshaft, in the case where a two-axis type balancer device is provided.

According to the inventions of claims 2 and 5, since the first and second balancer weights are arranged to be close to and outside first and second journal bearings bearing both sides of that portion of the crankshaft, which is coupled to the conrod, a dimension in the direction along the crankshaft can be minimized, so that in the case where a two-axis type balancer device is provided, it is possible to avoid a problem that enlargement around the crankcase in the direction along the crankshaft is caused due to arrangement of balancer weights away from that portion of a crankshaft, which is coupled to a conrod, in a direction along the crankshaft.

According to the invention of claim 3, since the first and second balancer shafts serve as coupling bolts for a left and right split type crankcase, it is possible to enhance a crankcase in coupling rigidity while suppressing complication of a construction and an increase in the number of parts, in the case where a two-axis type balancer device is adopted. Also, since the balancer shafts serve as crankcase coupling bolts, it is possible to reduce a space, in which exclusive crankcase coupling bolts are arranged, and in this point it is possible to avoid large-sizing around the crankcase.

According to the invention of claim 4, since the first and second balancer shafts are made close to the crankshaft so that loci of rotation of the first and second balancer weights partially overlap a locus of rotation of that portion of the crankshaft, which is coupled to a conrod when viewed in the direction along the crankshaft, it is possible to more surely avoid that large-sizing around the crankcase, which is caused due to arrangement of balancer shafts away from a crankshaft, in the case where a two-axis type balancer device is provided.

According to the invention of claim 5, since the balancer weights and the balancer driven gears are made integral with each other and rotatably supported on the balancer shafts, it suffices to rotatingly drive only the balancer weights and the balancer driven gears and it is unnecessary to rotatingly drive the balancer shafts themselves, according to which it is possible to achieve making effective use of an engine output. Also, freedom in assembly can be improved as compared with a constitution, in which balancer weights and balancer shafts are united together.

According to the invention of claim 6, since the crankcase-interior portions of the first and second balancer shafts are formed to have a smaller diameter than that of the weight supports, it is possible to make the balancer shaft further close to the crankshaft, thus enabling more surely avoiding large-sizing around the crankcase.

According to the invention of claim 7, since oil introduction passages are formed on the first and second balancer shafts to communicate the crank chamber and balancer-weight bearing portions to each other, that pulsation of pressure in the crank chamber, which is caused by up and down movements of the piston, causes the lubricating oil to be surely fed to the balancer weight bearing portions.

According to the invention of claim 8, since axes of rotation of the drive gears are made eccentric from axes of the balancer shafts, a simple construction and a simple operation of rotating the balancer shafts make it possible to regulate backlash relating to the balancer gears and the drive gears on the crankshaft, thus enabling preventing generation of noise.

According to the invention of claim 9, since both backlash regulation mechanisms for the first and second balancer shafts are provided on the same side of the engine, a work for regulating backlash on the both balancer shafts can be regulated on one side of the engine, and the work is made easy.

According to the invention of claim 10, since the drive gears on a side of the crankshaft to mesh with the balancer gears comprise a ring-shaped gear portion arranged around a disk-shaped base portion fixed to the crankshaft to be rotatable relative thereto and having teeth to mesh with the balancer gears, and U-shaped buffer springs interposed between the gear portion and the base portion, such compact construction absorbs shocks caused by torque nonuniformity of the engine or the like to be able to smoothly actuate the balancer device. 

1. A balancer device for engines, in which first and second balancer shafts are arranged in parallel to a crankshaft and the crankshaft rotatingly drives first and second balancer weights at the same speed as that of the crankshaft, characterized in that the, first balancer weight is arranged on one end of the first balancer shaft in a direction along the crankshaft, the second balancer weight is arranged on the other end of the second balancer shaft in the direction along the crankshaft, and the first and second balancer shafts are made close to the crankshaft so that loci of rotation of the first and second balancer weights partially overlap a locus of rotation of that portion of the crankshaft, which is coupled to a conrod when viewed in the direction along the crankshaft, wherein both sides of that portion of the crankshaft, which is coupled to the conrod, are born through first and second journal bearings by side walls of a crank chamber, the first and second balancer weights are rotatable supported by the first and second balancer shafts, and balancer driven sears formed on the first and second balancer weights mesh with crank drive gears, which are arranged close to and outside the first and second journal bearings.
 2. A balancer device for engines, in which first and second balancer shafts are arranged in parallel to a crankshaft and the crankshaft rotatingly drives first and second balancer weights at the same speed as that of the crankshaft, characterized in that the first and second balancer shafts serve as coupling bolts for a left and right split type crankcase, and the first and second balancer weights are rotatably supported by the first and second balancer shafts.
 3. The balancer device for engines, according to claim 2, wherein the first balancer weight is arranged on one end of the first balancer shaft in a direction along the crankshaft, the second balancer weight is arranged on the other end of the second balancer shaft in the direction along the crankshaft, and the first and second balancer shafts are made close to the crankshaft so that loci of rotation of the first and second balancer weights partially overlap a locus of rotation of that portion of the crankshaft, which is coupled to a conrod when viewed in the direction along the crankshaft.
 4. The balancer device for engines, according to claim 3, wherein both sides of that portion of the crankshaft, which is coupled to the conrod, are born through first and second journal bearings by side walls of a crank chamber, and balancer driven gears formed on the first and second balancer weights mesh with crank drive gears, which are arranged close to and outside the first and second journal bearings.
 5. The balancer device for engines, according to claim 4, wherein the first and second balancer shafts comprise weight supports for supporting the first and second balancer weights, and crankcase-interior portions arranged adjacent to the crankshaft, the crankcase-interior portions being formed to have a smaller diameter than that of the weight supports.
 6. The balancer device for engines, according to claim 4 or 5, wherein oil introduction passages are informed on the first and second balancer shafts to communicate the crank chamber and balancer-weight bearing portions to each other.
 7. The balancer device for engines, according to claim 6, wherein axes of rotation of the first and second balancer driven gears are made eccentric from axes of the first and second balancer shafts, and first and second balancer shafts are rotated to thereby enable regulating backlash between the first and second balancer driven gears and the first and second crank drive gears.
 8. The balancer device for engines, according to claim 7, further comprising backlash regulation mechanisms provided on the same end sides of the first and second balancer shafts in the direction along the crankshaft.
 9. The balancer device for engines, according to claim 8, wherein the first and second crank drive gears comprise a ring-shaped gear portion arranged around a disk-shaped base portion fixed to the crankshaft to be rotatable relative thereto and having teeth to mesh with the balancer driven gears, and U-shaped buffer springs interposed between the gear portion and the base portion. 